Refrigeration system, heat recovery system, refrigerated gas compression system and brayton cycle system

ABSTRACT

A refrigeration system in which cold refrigerant liquid is produced by permitting refrigerant vapor to flash from a batch of ambient temperature refrigerant liquid at progressively decreasing temperatures and pressures, which cools down the unvaporized liquid. Vapors are compressed by refrigerant compressors, cooled and condensed at ambient temperature, and used for succeeding refrigerant liquid batches. The above refrigerating system is preferably driven by a hot refrigerant liquid at progressively decreasing temperatures and pressures, thus cooling the remaining liquid to about ambient temperature. The vapors flashed drive refrigerant turbines, each of which drives a refrigerant compressor, and after expansion the vapor is cooled and condensed at ambient temperature. At the start of each cycle, as the batch of hot refrigerant liquid starts to flash the pressure available to the refrigerant expanders is at a maximum. At the same time, as the batch of ambient temperature refrigerant liquid starts to flash, the compression ratio required of the refrigerant compressors is at a minimum. Conversely, at the end of each cycle the pressure available to the expanders is at a minimum and the compression ratio required of the compressors is at a maximum. The refrigerant expanders and compressors therefore are valved sequentially so that at the start of the cycle the expanders are in series and the compressors are in parallel and at the end of the cycle the expanders are in parallel and the compressors are in series. In another form of the invention, in an open recuperated Brayton cycle engine, air is compressed with refrigeration at compressor inlet at least 50* F. below ambient and to from 0.9 to 1.1 of a calculated optimum compression ratio raised to the 0.286 power.

United States Patent Nebgen [54] REFRIGERATION SYSTEM, HEAT RECOVERYSYSTEM, REFRIGERATED GAS COMPRESSION SYSTEM AND BRAYTON CYCLE SYSTEMPrimary Examiner-William F. O'Dea Assistant E.raminer P. D. FergusonArlurney-Robert Ames Norton and Saul Leitner 57; ABSTRACT Arefrigeration system in which cold refrigerant liquid is produced bypermitting refrigerant vapor to flash from a batch BRAY TON CYCLE AIRCOMPRESSOR 1 June 13, 1972 of ambient temperature refrigerant liquid atprogressively decreasing temperatures and pressures, which cools downthe unvaporizcd liquid. Vapors are compressed by refrigerantcompressors, cooled and condensed at ambient temperature, and used forsucceeding refrigerant liquid batches.

The above refrigerating system is preferably driven by a hot refrigerantliquid at progressively decreasing temperatures and pressures, thuscooling the remaining liquid to about ambient temperature. The vaporsflashed drive refrigerant turbines, each of which drives a refrigerantcompressor, and after expansion the vapor is cooled and condensed atambient temperature.

At the start ofeach cycle, as the batch of hot refrigerant liquid startsto flash the pressure available to the refrigerant expanders is at amaximum. At the same time, as the batch of ambient temperaturerefrigerant liquid starts to flash, the compression ratio required ofthe refrigerant compressors is at a minimum. Conversely, at the end ofeach cycle the pressure available to the expanders is at a minimum andthe compression ratio required of the compressors is at a maximum. Therefrigerant expanders and compressors therefore are valved sequentiallyso that at the start of the cycle the expanders are in series and thecompressors are in parallel and at the end of the cycle the expandersare in parallel and the compressors are in series.

In another form of the invention, in an open recuperated Brayton cycleengine, air is compressed with refrigeration at compressor inlet atleast 50 F. below ambient and to from 0.9 to 1.1 ofa calculated optimumcompression ratio raised to the 0.286 power.

3 Claims. 1 Drawing Figure BRAYTON CYCLE EXPANDER GENERATOR couausnouCOOLER fipecu emronm FUEL V in, "H REFRIGERANT r HEATER 1 INTAKE AIR or,EXHAUST 22 h i a 2 9 2a 27) ls:2 is 12:2

7 n r: V r: '4

25 r T was...

34 REFRIG REFRIG TANK l compmzssoa EXPANDER REFRIG couueusen 4 COLD 50HOT 2 FLASH FLASH TANK TANK 1 s5 AMBIENT L. 36 STORAGE I cow TANK v 1STORAGE TANK 2 Z 5 PATENTEDJun z 3 i972 BRAYTON CYCLE BRAYTON CYCLE AIRCOMPRESSOR EXPANDER \k l GENERATOR T AT S f REc: AIR TOO COOLERTRECUPERATOR wgg gg --FuE| REFRIGERANT HEATER ATH k EXHAUST I2 HOTSTORAGE! 1 REFRIG REFRIG TANK COMPRESSOR EXPANDER I REFRIG CONDENSER 4HOT L FLASH i0 TANK AMBIENT e INVENTOR WILLIAM NEBGEN ATTORNEYREFRIGERATION SYSTEM, HEAT RECOVERY SYSTEM, REFRIGERATED GAS COMPRESSIONSYSTEM AND BRAYTON CYCLE SYSTEM BACKGROUND OF THE INVENTION Theefficiency of a refrigeration system is determined by the work which isrequired to remove the necessary quantity of heat from a process streamwhich it is desired to cool to some chosen temperature. When this heatis removed in stages of progressively lower temperatures, the work whichis required is reduced in accordance with the number of stages employed;the greater the number of stages, the less is the work required and thegreater is the efficiency of the refrigeration system.

The efficiency of a heat recovery system is determined by the work whichit produces from the heat which it removes from a process stream when itcools the process stream through a chosen temperature range. When thisheat is recovered and is used in stages of progressively lowertemperatures, the work which is produced is increased in accordance withthe number of stages employed; the greater the number of stages, thegreater is the work produced, and the greater is the efficiency of theheat recovery system.

If the gas which is about to enter a gas compressor is cooled byrefrigeration, less work is required to compress the gas through adesired compression ratio, and if the inlet gas is cooled to'a suitabletemperature the gas compression work which is saved is more than therefrigeration work which is needed to cool the gas. The total work,which is the sum of the cooled-suction gas compression work and therefrigeration work, is less than the uncooled-suction gas compressionwork alone. The amount of work which is saved depends on the efficiency(the number of stages) of the refrigeration system and on the gascompression ratio; the more efficient the system and the greater theratio, the greater the saving. With a given compression ratio, and whencooling to a given temperature, a two stage refrigeration systemprovides a greater saving than a single stage, and a three stage systemprovides a still greater saving.

In a Brayton cycle engine, air enters the engine at atmosphericpressure, is compressed, is heated and then is expanded back toatmospheric pressure. The net work output of the engine is therelatively small difference between two quite large numbers, i.e., it isthe difference in the total work produced by its air expander and thework consumed by its air compressor. The work produced by the airexpander of a 5.4 ratio simple Brayton cycle engine is about 2.77 timesthe net work output of the engine, and when the compressor takes suctionat ambient temperature (for example 100 F.) the work consumed by the aircompressor is about 1.77 times the net work output. If the ambienttemperature air is refrigerated before it enters the compressor, thework output of this Brayton cycle engine increases because thecompression ratio increases, and the expansion ratio increasesaccordingly; because the work produced by the air expander thereforeincreases; and also because the mass flow air through the engineincreases, due to the greater density of the cold air. The work which isrequired to refrigerate the inlet air must, of course, be deducted fromthe work which is produced by the Brayton cycle engine, but even when aninefficient single stage refrigeration system is used the refrigeratedsuction engine delivers more usable shaft work than does the same engineif it takes suction at 100 F.

SUMMARY OF THE INVENTION One aspect of the present invention concerns animproved refrigeration system.

In the refrigeration system, cold liquid phase refrigerant is usedcountercurrently to cool a process stream from about ambient temperatureto some chosen lower temperature. The cold refrigerant liquid thereby isheated to about ambient temperature. The liquid refrigerant ismaintained under sufficient pressure so that it does not boil at the toptemperature which it reaches as it countercurrently cools the processstream. A

batch of cold liquid refrigerant is produced by permitting a batch ofambient temperature refrigerant liquid to flash in a flash tank. Duringits flashing, the refrigerant liquid boils at progressively lowerpressures and temperatures. The vapor which evolves is removedcontinuously by refrigerant compressors and then is compressed, iscooled, and is condensed at about ambient temperature. The flashingresults in cooling down that portion of the refrigerant liquid whichdoes not flash, and when the liquid thus has been cooled downsuffciently it is stored temporarily in a cold storage tank andsubsequently it is used countercurrently to cool the process stream.Although the cold liquid is produced in batches, the cold storage tankspermits using it at a continuous rate to cool the process stream. Sincethe batch of ambient temperature refrigerant liquid is flashing atprogressively lower temperatures, heat is being removed from itcontinuously at progressively lower temperatures. The process thus isequivalent to one in which heat is removed and rejected in a very large(in theory, an infinite) number of stages of refrigeration.Consequently, in theory this refrigeration system is reversiblethermodynamically and is potentially the most efficient one which can beemployed.

The heat recovery system ambient temperature refrigerant liquid isheated and is maintained under sufficient pressure so that it does notboil while it is being heated. The hot liquid is stored temporarily in ahot storage tank. A batch of the heated liquid is transferred from thehot storage tank to a hot flash tank, where it is permitted to flash atdecreasing temperatures and pressures until finally it reaches apredetermined lower temperature, usually about ambient temperature. Theflash vapor passes through refrigerant expanders, each one driving acorresponding refrigerant compressor, as has been referred to above. Therefrigerant expander may drive its refrigerant compressor directly or itmay drive the compressor through a constant or through a variable speedchanging device. The flash vapor which leaves the last expander iscondensed at about ambient temperature. Since the batch of hotrefrigerant liquid is flashing at progressively lower temperatures, itis giving up heat continuously at progressively lower temperatures. Theprocess thus is equivalent to one in which heat is recovered and used ina very large (in theory, an infinite) number of stages of heat recovery.Consequently, in theory this heat recovery system is reversiblethermodynamically and is potentially the most efficient one which can beemployed.

At the beginning of the cycle the vapor coming off the hot flash tank isat its maximum temperature and pressure and can produce the maximumamount of expansion work. At the beginning of the cycle the vapor comingoff the cold flash tank is also at its maximum temperature and pressureand it requires the minimum amount of compression work. Therefore, atthe beginning of the cycle the available overall expansion ratio is usedin three expanders which are valved so that they operate in series, andthe required compression ratio is provided by the paired threecompressors which are valved so that they operate in parallel. (Threepairs of expanders and compressors form a very satisfactory compromisebetween efficiency and equipment costs, and so the general summary ofthe invention will be described in connection with such an arrangement,it being realized that there may be only two pairs of expanders andcompressors or more than three.) As the cycle progresses the pressure ofthe flashing refrigerant in the hot flash tank decreases and thepressure of the flashing refrigerant in the cold flash tank alsodecreases, so the work which is produced by each refrigerant expanderdecreases and the work which is required by each refrigerant compressorincreases. When the cycle has progressed to the point where eachexpander is unable to drive its paired compressor, the valving isswitched so that the hot vapor passes through two expanders in seriesand the cold vapor passes through two compressors in parallel. (Oneexpander and its paired compressor are cut off temporarily by valves.)The smaller overall expansion pressure ratio available at this time isused in only two, rather than three, expanders and each expanderproduces more work (SUlfiClEIlI to drive its paired compressor). After afurther drop in the pressure of the vapor from the hot flash tank,valves are again switched so that two expanders in parallel drive twocompressors in parallel. After a still further drop in the pressure ofthe hot flash tank, two expanders in parallel drive two compressors inseries, and finally, at the end of the cycle three expanders in paralleldrive three compressors in series.

By switching the compressor valves it is possible to operate with firstone, then two, and then three stages of compression. When the overallcompression ratio which is needed is produced in three stages ratherthan in one, less work is required by each compressor, but there is aconcomitant decrease in the total rate at which vapor flows from thecold flash tank, and there is a corresponding decrease in the rate atwhich the flash tank cools down.

By switching the expander valves it is possible to operate with firstthree stages, then two stages, and then one stage of expansion. When theoverall expansion ratio which is available is used in one stage ofexpansion rather than in three, more work is produced by each expander,but there is a concomitant increase in the total rate at which vaporflows from the hot flash tank, and there is a corresponding increase inthe rate at which the flash tank cools down.

At any given time during the course of the flash cycle the rate at whichvapor flows from the hot flash tank through the expanders is usuallyquite different from the rate at which vapor flows from the cold flashtank through the compressors, but at each and every instant of the flashcycle the work which is produced by each refrigerant expander is exactlythe same as the work which is required by its paired compressor.

During the whole of the cycle the vapor leaving the last refrigerantexpander is cooled to substantially ambient temperature and is condensedto a liquid. The refrigerant compressor discharge vapor likewise iscooled and condensed. The ambient temperature condensed refrigerantliquid may be the feed both for the cold flash tank and for the wasteheat recovery system. (it is, of course, possible, and in some cases itmay be desirable, to use different fluids for the refrigeration cycleand for the work producing cycle. in this case two separate condensersare used. When n cycle is completed the ambient temperature liquid whichis left in the hot flash tank is transferred to the ambient storagetank; the cold liquid which is left in the cold flash tank istransferred to the cold storage tank. Then a fresh batch of hotrefrigerant liquid is transferred from the hot storage tank to the hotflash tank, a fresh batch of ambient temperature refrigerant liquid istransferred from the ambient storage tank to the cold flash tank, andthe cycle is repeated.

In the refrigeration system of the present invention the heat value ofthe refrigeration work (W which is required to cool a process stream(for example, a lb. mol of gas, typically air) from ambient temperature(T,,) to some chosen lower temperature (T is ER [T(' in where C P themolal specific heat of the gas (for air about 7.0); T the chosencondensing temperature; AT the chosen temperature difference between thegas which is leaving the cooler and the refrigerant liquid which isentering the cooler; and E the chosen efficiency of the refrigerantcompressor.

As an example, assume that T, 560' R., T 580 R., AT E 0.8, and that itis desired to cool the gas to 400 R.,

]= 381 BTU/lb. mol.

W" C,-(T.| m' and for the assumed set of conditions,

381 O Q1 Q 0.34 BTU of work required for each BTU which is removed incooling the gas. Although for the chosen system C is dependent to someextent on the values of C T AT T and 15,, it is strongly dependent on TAs an illustration, when the values of C T AT T and E, are the same asin the previous example, but T is 450 R. rather than 400 R., C, =0.254rather than 0.34.

In the heat recovery system of the present invention, the heat value ofthe recovery work (W which is produced from the heat released by aprocess stream (for example a poundmol of gas, typically air) as thestream cools from a super-ambient temperature (T,,) to ambienttemperature T is u =CPEH [(Tu AT) T1 n w +1) where E is the chosenefficiency of the refrigerant expander, and AT is the chosen temperaturedifference between the gas and the refrigerant.

As an example, assume that E 0.8, AT 50, T 5 R., and that it isnecessary that the system produce 381 BTU of work per lb. mol,

from which T 950 R.

For any chosen heat recovery system a coefficient of performance C (at aparticular T can be calculated, where C is the ratio of W (in heatunits) to the heat which becomes available when a lb. mol of gas coolsfrom an initial temperature T to ambient temperature T Mathematically,

Wu c,- T,, T, y

and for the assumed set of conditions,

38] Cu MIOGSO 0. 14 ET U of work which is produced from each BTU ofavailable heat. For the chosen system, C H is somewhat dependent on thevalues of C E AT T,, and T but it is strongly dependent on T As anillustration, when the values of C E AT T,, and T are the same as in theprevious example, but T is 900 rather than 950 R., C 0.114 rather than0.14.

In another aspect of the present invention the ambient temperature powerproducing refrigerant liquid is used countercurrently to cool the gaswhich is discharged from a gas compressor, and the ambient temperaturepower producing refrigerant liquid is heated thereby. When the gassuction temperature and the gas compression ratio are suitably matched,the heat of compression of the gas heats the refrigerant power liquid toa temperature which is high enough so that the power liquid provides allof the work which is needed to refrigerate the gas which is about toenter the gas compressor, and no external work is needed to operate therefrigeration cycle. (Throughout the remainder of the specification thisselfdriven, sequentially valved refrigerant expander-compressor systemwill be referred to as the Treadwell System. This can be considered asthe most economical or the preferred way of operating this form of thepresent invention, although as an alternate the refrigeration cycleportion of the system may be powered in whole or in part by anindependent motor or steam turbine driver.) When the heat of gascompression provides the work of refrigeration, T is equal to the gascompressor discharge temperature, and T,,, (T /E o l) T wherein E is theefficiency of the gas compressor; r is the gas compression ratio; and nis the numerical value of adiabatic exponent (k-l/k) (for air, k L4, andn 0.286). When for a desired compression ratio r it is desired todetermine the matching T a trail T is selected and a corresponding T iscalculated from the preceding fonnula. The refrigeration work W, whichis required for a chosen refrigeration system to cool the air to thetrial T is calculated by the method previously explained. This W,, iscompared to the calculated heat recovery work W which is produced by achosen heat recover system (using the calculated T which corresponds tothe trial T A series of values of T is tried until the refrigerationwork which is required for the trial T is equal to the heat recoverywork which is produced when the corresponding calculated T is used.

As an example, assume that a compression ratio of 15.0 is desired andthat the self-driven Treadwell system is to be used. Several values of Ts are tried, which finally converge on 400 R., and as a check, thisvalue for T together with the desired value of 15.0 for r is substitutedin the previously given equation TH s/ c) '0" sln substituting,

T (400/l 85) (15.0 I) 400, from which T 950 R. It was previously shownthat with the Treadwell System, when 7",, 950 R., the work which isproduced by the heat recovery system supplies the work which is requiredby the refrigeration system when T 400 R.

With this suction temperature the gas compression requires only 71.5percent of the single stage adiabatic work which is required whensuction is taken at 560 R. in the prior art a compression ratio of 15.0cannot be achieved in a single stage compressor with ambient temperature(560 R.) suction because the discharge temperature of 1,330 R. (870 F.)is much too high, and because far too much compression work is consumed,so a compression ratio of this magnitude usually requires two expensive,intercooled stages of compression. However, when the Treadwell System isused to cool the suction gas to 400 R., the same compression ratio of15.0 is readily conducted in a single stage compressor which produces adischarge temperature of only 950 R. (490 F.), and at the same time thenet work is less than the work which is required by the more expensivetwo stage compressor. if other less efficient refrigeration and heatrecovery systems are used in place of the Treadwell System, morerefrigeration work is needed to cool the suction to 400 R., and lessheat recovery work is produced from the T of 950 R. Therefore, the heatrecovered cannot provide refrigeration to a temperature as low as 400 R,and the net gas compression work is greater.

When the Treadwell System is used to cool the gas which is about toenter a gas compressor, the subsequent work of adiabatic compressionclosely approximates the work of isothermal compression when theisothermal compression process is conducted at ambient temperature. infact, if A7}, and A7}, are made infinitely small, and T is made the sameas T,, when the Treadwell System is used to cool the suction gas thework of adiabatic compression exactly equals that of ambient temperatureisothermal compression.

Isothermal compression requires the least amount of work because intheory the process is reversible thermodynamically. With adiabaticcompression, the gas which is discharged from the compressor is at ahigher temperature than the gas which enters the compressor, and theheat energy which is required to produce this increase in temperature isprovided at the expense of additional work energy which has beendelivered to the compressor. The compressed gas is discharged from thecompressor at a relatively low temperature level and its heat normallyis wasted by being rejected to cooling water in an inter or an aftercooler. The direct rejection of this heat to cooling water is acompletely irreversible process thermodynamically. By contrast, in theheat recovery portion of the Treadwell System heat is also rejected tocooling water, but only after it has produced work in the refrigerantexpander. As a result, in the Treadwell System, in theory the heatrejection is completely reversible thermodynamically. Similarly, in therefrigeration portion of the Treadwell System, in theory the heatrejection is completely reversible. When in theory the Treadwell Systemis used with an adiabatic gas compressor, the gas initially is atambient temperature and after compression and heat recovery is also atambient temperature; the refrigeration process is reversible; the heatrecovery process is reversible; and the adiabatic compression process isreversible. Since the final temperature of the compressed gas is thesame as its initial temperature, and since in theory all of theprocesses involved are reversible, in theory adiabatic compression usingthe Treadwell System is equivalent to isothermal compression.

It will be noted that when the compressed gas supplies the heat whichfurnishes the work which is required by the refrigeration system, thesystem is self-regulating. if the gas compressor discharge temperaturerises, more heat is available, more work is developed and morerefrigeration work is available to lower the temperature of the gaswhich is about to enter the compressor. When this temperature islowered, the temperature of the gas which is discharged from thecompressor is in turn lowered. If the gas compressor dischargetemperature falls, less heat is available, less work is developed andless refrigeration work is available, so there is an increase in thetemperature of the gas which is about to enter the compressor, and thisincrease in turn raises the temperature of the gas which is dischargedfrom the compressor. This automatic self-regulation is an importantoperating advantage of this aspect of the present invention.

The refrigeration system also can be used to cool substances other thangas. In such a case heat from another source may be used to raise thetemperature of the refrigerant power liquid to a level high enough sothat it will provide all the work which is needed by the refrigerantcompressors. However, work is saved to the extent that waste heat isfurnishing at least some of the work for the refrigeration system, eventhough it may not be all of the work.

When the Treadwell System is used to cool the gas entering a gascompressor, the heat of gas compression need not be the only source ofheat for the power producing refrigerant liquid. There may be othersources, which further can increase the amount of self-drivenrefrigeration that can be produced, and this can permit a still lowergas compressor inlet temperature, with a still further saving incompressor work.

The combination of the Treadwell System with a Brayton cycle engineconstitutes another preferred form of the present invention wherein theair which is about to enter the compressor of a recuperated Braytoncycle engine is refrigerated and all the work of refrigeration isprovided by the heat which is recovered from the exhaust air which isleaving the recuperator of the same engine. According to the presentinvention, it has been discovered that the maximum Brayton cycle work isproduced when r has an optimum value defined by optimum r [(E E E T IT H(Eq. l), where E, X (r /r (the expansion ratio r, is smaller than thecompression ratio r because of parasitic pressure losses in the system);E and n are as previously defined; E is the air expander efficiency; Tis the air expander inlet temperature; and T is a chosen suctiontemperature, which is usually selected for practical reasons, such asthe cost and the performance of available refrigeration equipment.

It has been discovered according to the present invention that if theimprovement in performance is to be of practical significance, T shouldbe at least about 50 F. below the ambient air temperature that isordinarily encountered. It has further been discovered that satisfactoryresults can be achieved over a range of from 10 percent greater to 10percent smaller than the r actually calculated.

For any chosen T there is a unique value of r at which the work producedin a Brayton cycle is a maximum. The net work output, i.e. the Braytoncycle work less the refrigeration work, depends, of course, on theefficiency of the refrigeration system which is chosen, but once therefrigeration system is chosen, at the chosen T the net work output is amaximum at the same unique value of r at which (for the same T theBrayton cycle work output is a maximum.

It was shown previously that when the Treadwell System is used at theassumed conditions, the exhaust air must enter the heat recovery systemat 950 R. in order for the heat recovery system to provide therefrigeration work which is needed when the refrigeration system coolsfrom 560 R. to 400 R. the air which is about to enter the aircompressor. It was also shown previously that for maximum Brayton cyclework output the optimum r (E ETEpT /T (Eq. I).

As an example, assume that E 0.85; E 0.87; E,, (1.05)"= 1.014; T 1,960R. (1,500 R); and T =400 R.,

(0.85 X 0.87 X 1.014 X 1,960/400)"'*'= 1.915 r With this r E and T theair compressor discharge temperature (T is about 830 R. The recuperatortemperature approach (AT is the difference between the temperature ofthe compressed air which is entering the recuperator and the temperatureof the exhaust air which is leaving the recuperator and is, therefore,120 F. (950 830). This is approximately the difference between thetemperature of the compressed air which is leaving the recuperator andthe temperature of the exhaust air which is leaving the expander and isentering the recuperator. For the assumed r the assumed E and theassumed E,., the exhaust air leaves the expander at a temperature ofabout 1,160 R., so the compressed air is heated recuperatively to about1,040 R. (1,160 120) and the cycle operates at a thermal efficiency ofabout 40.2 percent.

lt is to be noted that when the Brayton cycle operating conditions (Tand its related r T E E and the Treadwell System operating conditions (TT AT AT E E are known, the recuperator temperature approach (AT isuniquely fixed. (This is true only when the work of refrigeration isprovided by the heat which is available in the recuperator exhaust.)Because the air expander inlet temperature (T the system pressure losseswhich fix E,,; the component efficiencies (E E E E the heat exchangertemperature approaches (AT AT the condensing temperature (T and theambient temperature (T are all constants for any selected system, therecuperator temperature approach (A- T is a function only of T and rSince T and r are related by the previously given (Eq. 1 for optimum rfor maximum Brayton cycle work output, r is a function only of ATTherefore, when for economic or other reasons a specific recuperatortemperature approach is chosen, this choice also determines the optimumr which is required for maximum work output.

This optimum r is given by the equation, (Eq. 2), r

2 (11. {1+ A THE.) CH

wherein E Ep, T T and E are as previously defined; AT is the chosenrecuperator temperature approach; C is the coefficient of performance ofthe chosen refrigeration system at the T at which it operates; and C isthe coefiicient of performance of the chosen heat recovery system at theT at which it operates. It is to be noted that although T does notappear explicitly in the equation, it is inherent in the calculation ofC and C The preceding equation holds true only when the work ofrefrigeration is supplied by the heat which is available in therecuperator exhaust.

In using this equation, a trial T is selected, and from r (E E E T /T Wrelating r to T at trial optimum r is calculated. The corresponding aircompressor discharge temperature T then is calculated and to it is addedthe desired recuperator temperature approach (AT to give 7' Thecoefiicient of performance C of the chosen heat recovery system iscalculated for this T and the coefficient of performance C R of thechosen refrigeration system is calculated for the same trial T Thevalues for T E AT C C T and E are substituted in Equation 2, and theresulting r is compared with the trial r Eq. 1. If this resulting r isnot the same as the trial r a new trial r is calculated from a new trialT a new C and a new C are calculated, and a new resulting r iscalculated. This procedure is repeated until the calculated resulting ris the same as the calculated trial r With a recuperator temperatureapproach of 10 F., (requiring an extremely large and very expensiverecuperator), by calculation the suction temperature is 442 R, theoptimum r is 1.825, the cycle thermal efficiency is 42.8 percent and thepower production is 7.7 BTU of work for each cu. ft. of air displaced bythe compressor, i.e. for each cu. ft. of compressor volumetric capacity.Other things being equal, the cost of a compressor is related to itsvolumetric capacity, and the cost of the compressor is a substantialpart of the cost of a Brayton cycle engine. The work produced per unitof compressor capacity is, therefore, a measure of the cost of theequipment used to produce power in a Brayton cycle engine.

With a recuperator temperature approach of F., the suction temperatureis 407 R., the optimum r is 1.898, the cycle thermal efficiency is 40.7percent and the power production is 8.6 BTU of work per cu. ft. ofcompressor capacity.

With a recuperator temperature approach of F., the suction temperatureis 400 F the optimum r5 is 1.915, the cycle thermal efficiency is 40.2percent, and the power production is 8.96 BTU of work per cu. ft. ofcompressor capacity.

With a recuperator temperature approach of F., the suction temperatureis 390 R., the optimum r is 1.94, the cycle thermal efficiency is 39.7percent, and the power production is 9.42 BTU of work per cu. ft. ofcompressor capacity.

When no recuperator is used, and the heat of the air which is leavingthe air expander is used only to power the refrigeration system, thesuction temperature is 343 R., the optimum r is 2.07, the cycle thermalefficiency is 37 percent, and the power production is 12.3 BTU of workper cu. ft. of compressor capacity.

When taking suction at 560 R., with an r of 1.62 and a recuperatortemperature approach of 150, in the prior art a recuperated uncooledBrayton cycle engine produces 394 BTU of work per cu. ft. of compressorcapacity, at a thermal efiiciency of about 29.2 percent. The recuperatorof this engine exhausts at about 1,120 R., and when this exhaust heat isused to make 50 psig steam in a waste heat boiler, the steam produces inan expensive separate steam turbine about 1.38 BTU of additional work,for a total of 5.32 BTU for each cu.ft. of capacity of the aircompressor of the Brayton cycle engine. The combined cycle thermalefficiency is about 39 percent.

With the same recuperator temperature approach of 150, an enginedesigned in accordance with the present invention has a suctiontemperature of 390 R., operates with a cycle thermal efficiency of about39.7 percent, and produces a net work output of about 9.42 BTU per cu.ft. of compressor capacity, which is about 2.39 times that of theuncooled standard recuperated cycle engine and about 1.77 times that ofthe uncooled combined recuperated cycle engine. It is to be noted thatthis 9.42 BTU work output is produced by the Brayton cycle engine alone,and that there is no need for an expensive separate power producingsteam turbine.

All the preceding Brayton cycle examples have been based on using theTreadwell System, but with a Brayton cycle it is possible to use other,less efficient combinations of self-driven refrigeration-heat-recoverysystems. For example, a single or a multi-stage refrigeration system canbe powered by a heat recovery system employing a single or a multi-stageboiler. For a chosen heat recovery system and for a chosen refrigerationsystem it is necessary to give due consideration to all the pertinentfactors, such as the heat exchanger temperature approaches, thecomponent efiiciencies, the number of stages, and the like, and tocalculate for each specific T its coefficient of performance C and foreach specific T its coeflicient of performance C The previouslydescribed procedure using a trail T is then employed to determine theoptimum for a chosen AT The preceding Eq. 2 for r is, therefore,applicable regardless of the type of refrigeration system or of the typeof heat recovery system which is used.

The thermal efficiency of a prior art Brayton cycle engine sufiersconsiderably if it becomes necessary to operate the engine at a reducedcapacity. For example, the previously described prior art engine, with arecuperator temperature approach of 150", an r of 1.62, and a suctiontemperature of 560 R. (ambient temperature 560 R.), operates at a designpoint thermal efficiency of 29.2 percent. When called on to operate at42 percent of its design point capacity, the same engine operates at athermal efficiency of l7.8 percent, which is only 61 percent of itsdesign point efficiency. This is because the most practical way toreduce the capacity of a prior art engine is to lower its firingtemperature, which simultaneously lowers its Carnot cycle efficiency aswell. In this example, at the design point the firing temperature is1,500 E, but at 42 percent capacity the firing temperature is only l,090F.

As engine built in accordance with the present invention, with arecuperator temperature approach of l50, an r of 1.94, and a suctiontemperature of 390 R. (ambient temperature 560 R.), operates at a designpoint thermal efficiency of 39.7 percent. When this engine is called onto operate at 42 percent of its design point capacity, it operates at athermal efficiency of 29.5 percent, which is almost 75 percent of itsdesign point efficiency. In this case, the capacity is reduced byallowing the level of suction refrigeration to rise to 560 R. The enginecapacity is easily varied between 42 percent and 100 percent of itsdesign capacity by adjusting the temperature level to which the suctionair is cooled. (This adjustment can readily be made in various ways, forexample, by throttling the flow of cooling water to the refrigerationcondenser.) When the temperature of the suction air is raised, theengine capacity is reduced. The firing temperature remains at 1,500 F.for this entire range of engine capacities. If the firing temperature islowered, the engine capacity can be reduced to even less than 42 percentof design capacity, at the penalty of somewhat lowered thermalefficiency.

BRIEF DESCRIPTION OF THE DRAWINGS The drawing shows, in purelydiagrammatic form, the combination of the refrigerant system for coolingthe air compressor for a recuperated open Brayton cycle.

DESCRIPTION OF THE PREFERRED EMBODIMENTS On the drawing, air at ambienttemperature T, enters the air cooler at the point marked Air Intake, andis cooled to a temperature T which is at least 50 F. below the ambienttemperature at which it is intended to operate the engine. The cooledair enters the air compressor, in which it is compressed through thecalculated optimum compression ratio r The compressed air enters therecuperator at a discharge temperature which is determined by thisoptimum r and suction temperature T In the recuperator the air is heatedby heat exchange with the exhaust from the Brayton cycle expander andpasses into a conventional Brayton cycle combustion chamber. In thischamber fuel is burned and the temperature of the compressed air israised further to T which is the maximum temperature that the materialsof the expander can withstand. The maximum permissible level of T,- isin no sense changed by the present invention.

The compressor is driven by the expander. The difference in the workwhich is produced by the expander and the work which is required by thecompressor constitutes the net work output of the Brayton cycle. This issymbolized on the drawing by the power output shaft being connected toand driving the generator.

The expander exhaust gases go to a recuperator, which they leave at atemperature T T is determined by the discharge temperature of thecompressed air and by the temperature differential AT indicated. Theexhaust gases then pass through a refrigerant heater in which pump (2)keeps the refrigerant liquid at a sufficient pressure so that it doesnot boil. The amount of liquid which goes to the heater is determined bythe adjustment of valves (5) and (6). In the heater the liquidrefrigerant is heated up to temperature T minus the small temperaturedifferential AT which is required for heat exchange. The exhaust gasesthen are exhausted as indicated, ordinarily at ambient temperature plusthe same small temperature differential AT The hot refrigerant liquidflows from the refrigerant heater into a suitably insulated hot storagetank I From time to time valve (4) is opened, and a batch of hot liquidis transferred from hot storage tank (1) to hot flash tank (3). Theliquid holding capacity of hot storage tank (I) is sufficiently greaterthan that of hot flash tank (3) to permit substantially continuousoperation. The drawing is diagrammatic, so only a single hot flash tankis shown, but multiple tanks can be used, if desired.

In hot flash tank (3) the heated refrigerant liquid, initially undersuch pressure as may be needed to prevent boiling in the refrigerantheater, flashes at decreasing temperatures and pressures until itreaches a minimum temperature and pressure, normally about ambienttemperature. Valve (10) then is opened, and the remaining unvaporizedliquid is permitted to flow into ambient storage tank l 1).

Three refrigerant expanders 7), (8) and (9) constitute the powergenerating portion of the refrigeration system. The pattern of flowthrough the expanders is controlled by valves (l2), (l3), (l4), (l5),(l6), (17), (18), and (19). At first, when the vapor in the hot flashtank is at maximum temperature and pressure, valves (l2), (14), (17) and(19) are opened, and valves (13},(15), (l6) and(l8) are closed. As aresult, refrigerant vapor passes in series through (7), (8) and 9).These expanders drive corresponding refrigeration compressors (21), (22)and (23). This is symbolized on the drawing as a common shaft connectingexpander (7) and compressor (23), a common shaft connecting expander (8)and compressor (22), and a common shaft connecting expander (9) andcompressor 2i At the start, the temperature and pressure in hot flashtank (3) is at a maximum and the flash vapor passes through expanders(7), (8) and (9) in series. At the same time, the pressure andtemperature in cold flash tank (30) is at a maximum, and the load onrefrigeration compressors (2! (22) and (23) is at a minimum. The patternof flow through these compressors is controlled by valves (20), (24),(25), (26), (27), (28), (29) and (33). At the start the threecompressors operate in parallel, valves (20), (24), (26), (27), (29) and(33) being open, and valves (25) and (28) being closed. The load oncompressors (21), (22) and (23) increases as the temperature andpressure of the refrigerant in cold flash tank (30) drops. Whenexpanders (7), (8) and (9) can no longer produce sufficient power todrive the compressors, valves (l7), (19), (20) and (26) are closed. Thishas the effect of cutting off expander (9) and compressor (21 and nowexpanders (7) and (8) in series drive compressors (22) and (23) inparallel.

After a further lapse of time, the pressure and temperature of therefrigerant in hot flash tank (3) and of the refrigerant in cold flashtank (30) drops. When the load on compressors (22) and (23) increasesand the power output of expanders (7) and (9) decreases to the pointwhere the expanders can no longer drive the compressors, valves l3) and(15) are opened and valve (14) is closed. Now expanders (7) and (8) inparallel drive compressors (22) and (23) in parallel.

After a further drop in the temperature and pressure of the refrigerantin tanks (3) and (30), valves (27) and (29) are closed and valve (28) isopened. This results in two expanders, (7) and (8), in parallel drivingtwo compressors, (22) and (23), in series.

When the temperatures and pressures in tanks (3) and (30) have droppedstill further, valves (16), (19), and are opened and valve (24) isclosed. Now the three expanders (7), (8) and (9) operate in parallel todrive the three compressors (21), (22) and (23) in series. It will benoted that during the whole operation exhaust vapors from the expandersand compressed vapors from the compressors flow into a conventionalwater cooled refrigeration condenser (34), where the vapors arecondensed at practically ambient temperature. The condensate isdischarged into ambient storage tank (11). When expanders (7), (8) and(9) no longer have sufficient power to drive compressors (21), (22) and(23), valve (10) is opened, and the unvaporized liquid in hot flash tank(3), now at substantially ambient temperature, also is discharged intoambient storage tank (11). The unvaporized cold liquid in cold flashtank is discharged into cold storage tank (31) through valve As in thecase of hot storage tank (1 cold storage tank (31) should havesufficient capacity so that continuous operation is possible.

In the meantime pump (32) continuously has been pumping cold refrigerantliquid from cold storage tank (31) through the air cooler, which hasbeen mentioned above. Flow of the cold liquid is controlled by valve(36). It will be seen from the drawing that the refrigerant liquidleaves the air cooler at substantially ambient temperature and flowsinto ambient storage tank (11), in which it is joined by the condensatewhich is formed in refrigeration condenser (34). Valves (l0) and (35)now are closed and valves (4) and (6) are opened. A new batch ofrefrigerant liquid from hot storage tank (1) thus is introduced into hotflash tank (3) and a new batch of ambient temperature liquid thus isintroduced into cold flash tank (30). The refrigeration cycle then isrepeated. The system is self-regulating. If temperature T at the inletof the air compressor tends to increase, the temperature of thecompressed air entering the recuperator also increases and so, likewise,does T This, in turn, heats the refrigerant liquid to a highertemperature. The flashing of this hotter liquid in hot flash tank (3)produces more power which in turn reduces the temperature of therefrigerant in cold storage tank (31) and lowers T If T tends todecrease, the process is reversed. This self-regulation is an advantagewhen the Treadwell System of refrigeration is combined with therecuperated Brayton cycle.

The preferred embodiment shown by the drawing utilizes all of theadvantages of a full Treadwell System and represents a preferredmodification, but the invention is not limited to using all of theadvantages, and may use only part of them.

It will be noted that combining the Treadwell System of refrigerationwith a Brayton cycle gives optimum results. However, the air which isabout to enter the compressor of the Brayton cycle may be cooled by anyother refrigerating means, though with some loss in overall work outputor efficiency.

Looking at the refrigeration system alone apart from a Brayton cycle,the parallel-series flow pattern of the refrigeration compressors may beused with any type of driver, and is not restricted to waste heatepowered expanders. However, where waste heat is available, it isdesirable to use it to the maximum extent possible. As has been pointedout before, in the Treadwell System it is not essential that the samerefrigerant which is used in the power producing cycle be used in therefrigeration cycle, but when the Treadwell System is combined with aBrayton cycle, it is ordinarily more convenient and economical to usethe same refrigerant liquid both for power production and refrigeration.

I claim:

l. A refrigerating system in which a volatile refrigerant is chilled byvolatilizing a portion thereof to produce cold liquid refrigerant, thevolatilization being at decreasing pressures, comprising in combinationa flash tank for liquid refrigerant, a plurality of compressors, powersources for driving each compressor, each compressor having a suctioninlet for a volatilized refrigerant and a compressed vapor outlet,valved conduits extending from the refrigerant flash tank at a levelabove liquid refrigerant therein to t e suction inlets of thecompressors, and valved conduits connecting the compressed vapor outletfrom at least one of the compressors to the suction inlet of anothercompressor, whereby upon actuation of the valves vapors from therefrigerant flash tank can be directed to the suctions of all of thecompressors in parallel or to two or more compressors in series, acooled compressed vapor condenser, a source of coolant to the condenserat a temperature sufficiently low to condense compressed refrigerantvapors in the condenser, valved conduits connecting the outlets of thecompressors to the cooled compressed vapor condenser, a condensedrefrigerant vapor storage container and conduit means connecting thecondenser thereto, whereby condensed refrigerant vapors flow into thestorage tank, means for controlling the valves in the valved conduits tooperate the compressors in parallel until the pressure in therefrigerant flash tank has reached a predetermined value and thensuccessively connecting compressors in series until finally all of thecompressors are in series and the pressure in the refrigerant flash tankreaches a minimum, means for then discharging the unvaporized and cooledrefrigerant liquid from the flash tank to a cold refrigerant liquidstorage tank, means for transferring a fresh charge of condensedcompressed refrigerant vapors from the condensed vapor storage tank andmeans for repeating the above cycle by connecting the flash tank to theinlets of the compressors in parallel.

2. A refrigerating system according to claim 1 in which each compressoris connected to its own power turbine to drive it, a source of hotliquid refrigerant, a hot liquid refrigerant flash tank, means fortransferring a charge of hot refrigerant liquid to the flash tank, eachpower turbine having a turbine vapor inlet and outlet, valved conduitsconnecting the turbine inlets to the hot flash tank at a point above thehot refrigerant liquid level therein, valved conduits connecting powerturbine outlets to power turbine inlets of another power turbine, meansfor actuating the valves in the conduits so that at first the powerturbines are in series while the refrigerant compressors are inparallel, whereby the hot refrigerant liquid flashes to predeterminedtemperature and pressure, then connecting power turbines in parallel ascompressors are con-nected in series, until finally all of the powerturbines are in parallel and all of the compressors are in series untilthe hot refrigerant flash tank has flashed to a minimum temperature andpressure, valved conduits connecting power turbine exhausts to thecompressed refrigerant vapor condenser, whereby these turbine exhaustsare condensed to a liquid and added to the condensed refrigerant liquidstorage tank, and finally, when the hot and cold refrigerant flash tankshave flashed to predetermined minimum temperatures and pressures, meansto transfer the unflashed refrigerant from the hot refrigerant flashtank to the compressed refrigerant vapor storage tank and to introducefrom the hot refrigerant storage tank to the hot refrigerant flash tanka fresh charge, and repeat the above cycle.

3. Refrigerating system according to claim 2 in which the number ofturbines and compressors is three each.

1. A refrigerating system in which a volatile refrigerant is chilled byvolatilizing a portion thereof to produce cold liquid refrigerant, thevolatilization being at decreasing pressures, comprising in combinationa flash tank for liquid refrigerant, a plurality of compressors, powersources for driving each compressor, each compressor having a suctioninlet for a volatilized refrigerant and a compressed vapor outlet,valved conduits extending from the refrigerant flash tank at a levelabove liquid refrigerant therein to the suction inlets of thecompressors, and valved conduits connecting the compressed vApor outletfrom at least one of the compressors to the suction inlet of anothercompressor, whereby upon actuation of the valves vapors from therefrigerant flash tank can be directed to the suctions of all of thecompressors in parallel or to two or more compressors in series, acooled compressed vapor condenser, a source of coolant to the condenserat a temperature sufficiently low to condense compressed refrigerantvapors in the condenser, valved conduits connecting the outlets of thecompressors to the cooled compressed vapor condenser, a condensedrefrigerant vapor storage container and conduit means connecting thecondenser thereto, whereby condensed refrigerant vapors flow into thestorage tank, means for controlling the valves in the valved conduits tooperate the compressors in parallel until the pressure in therefrigerant flash tank has reached a predetermined value and thensuccessively connecting compressors in series until finally all of thecompressors are in series and the pressure in the refrigerant flash tankreaches a minimum, means for then discharging the unvaporized and cooledrefrigerant liquid from the flash tank to a cold refrigerant liquidstorage tank, means for transferring a fresh charge of condensedcompressed refrigerant vapors from the condensed vapor storage tank andmeans for repeating the above cycle by connecting the flash tank to theinlets of the compressors in parallel.
 2. A refrigerating systemaccording to claim 1 in which each compressor is connected to its ownpower turbine to drive it, a source of hot liquid refrigerant, a hotliquid refrigerant flash tank, means for transferring a charge of hotrefrigerant liquid to the flash tank, each power turbine having aturbine vapor inlet and outlet, valved conduits connecting the turbineinlets to the hot flash tank at a point above the hot refrigerant liquidlevel therein, valved conduits connecting power turbine outlets to powerturbine inlets of another power turbine, means for actuating the valvesin the conduits so that at first the power turbines are in series whilethe refrigerant compressors are in parallel, whereby the hot refrigerantliquid flashes to predetermined temperature and pressure, thenconnecting power turbines in parallel as compressors are con-nected inseries, until finally all of the power turbines are in parallel and allof the compressors are in series until the hot refrigerant flash tankhas flashed to a minimum temperature and pressure, valved conduitsconnecting power turbine exhausts to the compressed refrigerant vaporcondenser, whereby these turbine exhausts are condensed to a liquid andadded to the condensed refrigerant liquid storage tank, and finally,when the hot and cold refrigerant flash tanks have flashed topredetermined minimum temperatures and pressures, means to transfer theunflashed refrigerant from the hot refrigerant flash tank to thecompressed refrigerant vapor storage tank and to introduce from the hotrefrigerant storage tank to the hot refrigerant flash tank a freshcharge, and repeat the above cycle.
 3. Refrigerating system according toclaim 2 in which the number of turbines and compressors is three each.